mud pump petrowiki factory

Within each Supply Base, you can usually find what "Mud" companies are working offshore from this base as they are represented in the form of a Liquid Mud Plant. This mud plant is used to manipulate (mix, dilute, treat, adjust) various drilling fluids to be used in wells supported by this Supply Base. You can also find expanded drilling waste management (DWM) processing here, typically in the form of centrifuges used to treat used mud that has been back loaded to the supply base. OSV (offshore supply vessels) are used to transport any drilling or completion fluids to and from the LMP that cannot be mixed up ort treated on location. The LMP is in place for one primary purpose and that is to, "Keeping good mud properties in the wellbore as much as possible".

There exists a few downsides to any LMP and that is the capability of installing the necessary infrastructure for any LMP. This infrastructure being a series of tanks with necessary missing hoppers, pumps & piping. Any tank format chosen for a particular site come with very heavy footprints so a very well compacted site need be present quayside.

Many remote operations do not allow or justify the creation of any LMP because of initial cost or the lack available land and / or dock space. If this is the case, an offshore drilling operation is then dependent upon long boat runs to the chosen Supply Base. OSV (boat) runs in excess of several days is not uncommon. This waiting period for a specific loaded OSV, can equate to extended rig down time. Enter the Liquid Mud Plant Barge to bring Supply base type support alongside the drilling operation.

Water depth, environmental demands, client drilling program mud requirements, contract length, availability all qualify any LMP Barge selection. For any barge, assignment will be in "sheltered waters" or as a rule of thumb in waters < 100m in depth. Mooring analysis will be done to determine single, 4 or 8 point mooring pattern.

There are three common types of gas engines used for beam pumping units: two-cycle, slow-speed engine; four-cycle, slow-speed engine; and four-cycle, high-speed engine. The characteristics of these engines are summarized here, and the detailed comparisons and field experiences have been published elsewhere.

The electric motor most commonly used for beam-pumping installations is an alternating-current (AC), three-phase, squirrel-cage induction motor. These motors are used for the following reasons:

A general guide of motor size vs. V is 115 or 230 V for single-phase motors; 115, 230, 460, or 575 V for polyphase motors up to 50 HP; and 460, 575, or 796 V for polyphase motors 50 to 200 HP. Motors for pumping units come in a variety of common sizes: 1, 1.5, 2, 3, 5, 7.5, 10, 15, 20, 25, 30, 40, 50, 60, 75, 100, and 125 HP.

Motors can be purchased in six standard synchronous speeds, with the 1,200-rpm motor being the most commonly used in oilwell pumping. Multiple-HP-rated motors that may be either dual- or triple-rated are sometimes used for oilwell pumping; the triple-rated is more common. Changing one of these motors from one HP rating to another requires changing leads in the motor housing, which in turn changes the motor"s internal wiring system. Any capacitors, fuses, or overload relays in the circuit will also require evaluation and possible revision at the same time to make sure it agrees with the new voltage/current requirements.

NEMA presents five general design standards that provide for varying combinations of starting current, starting torque, and slip. The most commonly recommended electric motor for pumping units is a 1,200-rpm NEMA Design D. It has a normal starting current, a high starting torque (272% or more of full-load torque), and a high slip (5 to 8%). Because Design D specifications are not drawn as closely as they are for other designs, manufacturers have developed several designs with variations in slip that still fall within Design D specifications.

A power factor determines the amount of line current drawn by the motor. A high power factor is desirable because it is important in reducing line losses and minimizing power costs. A lower power factor means that the unit is not operating as efficiently as it should. Oversized motors tend to have low power factors. Typically, a NEMA D has a power factor of 0.87 when fully loaded, but decreases to 0.76 at half load. Usually, units must operate at a power factor of greater than 0.80 to avoid penalties from the power companies; thus, optimization of the pumping unit"s size and motor needs to be considered as the well-fluid volume changes.

Using capacitors can increase power factors. To determine if and how much capacitance is needed, determine the power factor of an installation upon initial startup and then decide if a correction is justified. If a pumping-unit motor has a low power factor, a capacitor can be placed between the motor and disconnect. Because of the possibility of electrical shock, only qualified personnel should make this connection. Remember that changing producing conditions might require that the power factor be checked and that the motor-overload relays be resized if the capacitor is on the load side of the overload relays.

When a motor is used for a cyclic load, such as oilwell pumping, it will be thermally loaded more than the same average load applied on a steady-state basis. HP ratings of electrical motors depend on how much the temperature increases in the motor under load. A motor functioning cyclically must be derated from its full-load nameplate rating.

There are four basic types of motor enclosures: drip-proof guarded, splashproof guarded, totally enclosed fan cooled (TEFC), and explosion proof. "Guarded" refers to screens used over air intakes to prevent the entrance of rodents or other foreign items. The TEFC enclosure provides the maximum protection for the interior of the motor. The drip-proof motor should prove adequate for most pumping-unit installations in which the motor is elevated. This type of construction is built with a closed front-end bell to eliminate the entry of horizontal rain, sleet, or snow into the motor. The splashproof motor affords somewhat more protection against splashing liquids than does the drip-proof one. The preferred enclosure sets the motor on or close to the base; the explosion-proof enclosure will seldom be required. Motor-high mounts on pumping units have also been useful in protecting the motor from sand or snow.

Slip is the difference between motor synchronous speed and speed under load, usually expressed in percent of synchronous speed. Synchronous speed is the theoretical, no-load speed of the motor. Slip characteristics are very important because they will determine how much HP can be converted to torque to start the gearbox gears turning. A high-slip motor permits the kinetic energy of the system to assist in carrying the peak-torque demands. A low-slip motor will respond to the instantaneous demand; in other words, the high-slip motor slows down more under peak torque demands than the low-slip motor. The result is that the high-slip motor will require lower peak currents than the low-slip motor. How high the motor slip should be for pumping installations is debatable; however, Howell and Hogwood stated, "A slip greater than 7 to 8% offers no additional advantages from the overall pumping efficiency standpoint."

The electrical equipment must be properly grounded. Good grounding procedures are essential to personnel safety and good equipment operation. It is recommended that reference be made to the Natl. Electrical Code and the Natl. Electrical Safety Code to ensure safe grounding is met. Particular attention should be given to the connection of the ground wire to the well casing. The connection should be located where it will not be disturbed during well-servicing operations and should be mechanically secure. Periodic (yearly is recommended as a minimum) continuity measurements should be made with a volt-/ohmmeter between "a new clean spot" (not where the ground wire is terminated) on the well casing and new spot on each piece of grounded equipment. The resistance measured between any piece of equipment and the casing should not exceed 1 ohm. The resistance measured between the pumping-unit ground system and another nearby moisture ground should not exceed 5 Ω. However, these measurements should to be checked with current circulating through the system to determine if the ground is good.

There are seven HP values that should be considered in the proper design and operation of sucker-rod-pumped wells; these are hydraulic, friction, polished-rod, gear-reducer, V-belt drive, brake, and indicated.

Hydraulic HP (HHP) is the theoretical amount of work or power required to lift a quantity of fluid from a specified depth. This is a theoretical power requirement because it is assumed that there is no pump slippage and no gas breakout. The HHP, thus, is the minimum work expected to lift the fluid to the surface and can be found with the following equations:

Friction HP (FHP) is the amount of work required to overcome the rubbing-contact forces developed when trying to lift the fluid to the surface. This friction can be caused by a number of sources including plunger-on-barrel friction; rod- and/or coupling-on-tubing wear; sand, scale, and/or corrosion products hindering pump action, rods, and couplings moving through the fluid; fluid moving up the tubing; normal and excessive stuffing-box friction; and liquid and gas flowing through the flowline and battery facilities. FHP, thus, is dependent on factors such as how straight and deep the well is, the fluid viscosity, the pumping speed, and the tubing/rod buckling. In most situations, unless we know all of these factors, we do not know what FHP is. However, for design purposes, API RP11L calculations assume the friction effects, which show up in the peak and minimum polished-rod loads and in the calculation of polished-rod HP (PHP).

V-belt-drive HP (VHP) is the maximum power required by the V-belts to be transmitted to the gear reducer. API Spec. 1BVHP for a beam-pumping unit is as follows:

Prime movers—whether with a gas engine or an electric motor—run at a speed of 300 to 1,200 rpm. This speed must be reduced to the required pumping-unit speed of 2 to 25 spm. This is accomplished with sheaves, V-belt drives, and gear reducers. A sheave is a grooved pulley, and its primary purpose is to change the speed between the prime mover and the gearbox. The belt—usually a V-belt —is a flexible band connecting and passing around each of the two sheaves. Its purpose is to transmit power from the sheave on the prime mover to the sheave on the pumping unit. It is important to understand the basics of sheaves and V-belt to know how to select a sheave for a certain pumping speed and to determine the number of V-belt needed.

Sheaves come in different widths and have from 1 to 12 grooves. They are selected on the basis of the pitch diameter (PD) relative to how many spm the unit will pump. New beam-pumping units can be purchased with different-sized sheaves on the reducer. Sheaves can also be purchased to accept different V-belt cross sections. A pumping-unit sheave should be selected that will allow as much speed variation (up and down) from the design speed as is practical without violating API Spec. 1BVHP is shown in Eq. 11.15. Only the grooves closest to the prime mover and the gear reducer should be filled, and only enough belts to transmit the VHP should be installed because of the following considerations:

Pumping-unit manufacturers usually list all unit-sheave sizes in their catalogs. Motor sheaves are available with various PDs and numbers of belt grooves. Table A.1 in API Spec. 1B contains commonly available sheaves. Because of availability, motor sheaves should be selected from those listed in the top portion of the table.

A V-belt has a trapezoidal cross section that is made to run in sheaves with grooves that have a corresponding shape. It is the workhorse of the industry, available from virtually every V-belt distributor, and it is adaptable to practically any drive. It was designed to wedge in the pulley, thereby multiplying the frictional force produced by the tension; this, in turn, reduces the belt tension required for an equivalent torque. Remember, the purpose of the belt is to transmit power from the sheave on the prime mover to the sheave on the pumping unit. Therefore, the number and size of the belts needed depend on the amount of power to be transmitted.

The first step in designing the V-belt drive for a pumping unit consists of selecting a sheave for the unit and the prime mover. To do this, the desired pumping speed (N), along with the speed (in rpm) of the prime mover and gear ratio, must be known. If the other parameters are known, this equation can be rearranged to determine any required factor:

The largest motor sheave in this group will provide for the greatest reduction in pumping speed for future operations merely by changing motor sheaves.

A double-reduction unit run by an electric motor will require a speed reduction through the V-belt drive of approximately 2:1 at fast pumping speeds. At slow speeds, the ratio will be 6:1. When two belt sections are offered for the unit sheave, the smaller belt section will allow the use of a smaller motor sheave and a lower pumping speed. In most cases, the smaller belt section, with one of the two largest-unit sheaves, will offer the greatest flexibility.

A double-reduction unit run by a slow-speed gas engine will require a speed reduction of 1:1 at a fast pumping speed; at a slow pumping speed, the ratio will be 3:1. In these cases, speed reductions (which may be anticipated through the drive) should be checked with the proposed unit and prime mover. If little or no speed reduction will ever be required through the V-belt drive, one of the two smaller-unit sheaves will enable the use of a smaller (and less-expensive) prime-mover sheave. The larger belt section could also be used and may require fewer belts.

Given: gear-reducer sheaves available from the pumping-unit manufacturer"s catalog: 20-, 24-, 30-, 36-, and 38-in. PD-3C. Assume that the prime mover"s average rpm = 1,120. The smallest C-section motor sheave that should be considered = 9 in. PD (i.e., 9.4-in. OD in Table 3.1 of API Spec. 1B). The largest sheave that should be considered to keep the design PD velocity at less than 5,000 ft/min = 16-in. PD (calculations indicate a 17-in. PD, but page 32 of API Spec. 1B indicates that 17-in. PD C-section sheaves are not generally available; economics should discourage engineers and others from recommending sheaves not listed). The liquid to be pumped has a viscosity of approximately 1 cp. The pumping-unit gear ratio is 28.67. The maximum speed with an 86-in. stroke should result in an acceleration factor of 0.3, in which the maximum spm ≤ (0.3 × 70,500/86) 0.5 ≤ 15.7. The minimum speed with an 86-in. stroke should result in an acceleration factor ≤ 0.225, in which the minimum spm ≤ (0.225 × 70,500/86) 0.5 ≤ 13.6.

Solving for pumping speeds from Eq. 11.20 = [prime-mover speed (rpm) × prime-mover-sheave PD]/[(gear-reducer sheave PD) × (1/pumping-unit gear ratio)]. For example, 1,120 × 9/20 × 1/28.67 = 17.1. The rest of the speeds can be calculated similarly for the different available gear-reducer sheaves, and the smallest or largest prime-mover sheaves. The summary of these calculations is shown in Table 11.11.

The normal ESP system configuration is shown in Fig. 13.1. It shows a tubing-hung unit with the downhole components comprising of a multistage centrifugal pump with either an integral intake or separate, bolt-on intake; a seal-chamber section; and a three-phase induction motor, with or without a sensor package. The rest of the system includes a surface control package and a three-phase power cable running downhole to the motor. Because of the ESP’s unique application requirement in deep, relatively small-bore casings, the equipment designer and manufacturer are required to maximize the lift of the pump and the power output of the motor as a function of the diameter and length of the unit. Therefore, the equipment is typically long and slender. The components are manufactured in varying lengths up to approximately 30 ft, and for certain applications, either the pump, seal, or motor can be multiple components connected in series.

Throughout their history, ESP systems have been used to pump a variety of fluids. Normally, the production fluids are crude oil and brine, but they may be called on to handle liquid petroleum products; disposal or injection fluids; and fluids containing free gas, some solids or contaminates, and CO2 and H2S gases or treatment chemicals. ESP systems are also environmentally esthetic because only the surface power control equipment and power cable run from the controller to the wellhead are visible. The controller can be provided in a weatherproof, outdoor version or an indoor version for placement in a building or container. The control equipment can be located within the minimum recommended distance from the wellhead or, if necessary, up to several miles away. API RP11S3 provides the guidelines for the proper installation and handling of an ESP system. Table 13.1, some of which are discussed later in this chapter.

The ESP is a multistage centrifugal type. A cross section of a typical design is shown in Fig. 13.2. The pumps function is to add lift or transfer pressure to the fluid so that it will flow from the wellbore at the desired rate. It accomplishes this by imparting kinetic energy to the fluid by centrifugal force and then converting that to a potential energy in the form of pressure.

In order to optimize the lift and head that can be produced from various casing sizes, pumps are produced in several diameters for application in the most common casing sizes. Table 13.2 lists some common unit diameters, flow ranges, and typical casing sizes in which they fit.

Shaft. The shaft is connected to the seal-chamber section and motor by a spline coupling. It transmits the rotary motion from the motor to the impellers of the pump stage. The shaft and impellers are keyed, and the key transmits the torque load to the impeller. As was mentioned earlier, the diameter of the shaft is minimized as much as possible because of the restrictions placed on the pump outside diameter. Therefore, there are usually several shaft material options available, depending on the maximum horsepower (HP) load and corrosion protection required.

Housing. The housing is the pressure-containing skin for the pump. It holds and aligns all the components of the pump. There are several material options available for different application environments. For additional corrosion protection, there are several coatings that can be applied.

Several different styles of intakes can be selected. They allow for entrance of the fluid into the bottom of the pump and direct it into the first stage. Integral intakes can be threaded directly into the bottom of the housing during the manufacturing assembly process, while others are separate components, which are bolted on to the bottom pump flange.

A standard intake has intake ports that allow fluid to enter the pump. It is used when the fluid is all liquid or has a very low free-gas content. The intake shown in Fig 13.2 would be a standard intake if the reverse-flow screen were omitted.

A reverse-flow intake is used when the free-gas content in the fluid is high enough to cause pump-performance problems. The pump in Fig. 13.2 is shown with a reverse-flow design. The produced fluid with free gas flows up the outside of the reverse-flow intake screen, makes a 180° turn to enter through the perforations or holes at the top of the screen, flows back down to the intake ports and then back up to the first pump stage. These reversals in direction allow for a natural separation of the lighter gases from the liquid. The separated gas travels up the casing annulus and is vented at the wellhead. Another style is shown in the right-hand graphic of Fig. 13.3, which has a longer reversing path than does the intake with the screen.

The next step in handling free gas with an ESP involves downhole mechanical separation devices such as separator intakes. These devices take the fluid that enters its intake ports, impart a centrifugal force to it, vent the lighter-density fluid back to the annulus, and transfer the heavier-density fluid to the first pump stage. The heavier-density fluid, which is routed to the pump, has been either fully or partially degassed. Two of these devices are shown in the left-hand and center graphics of Fig. 13.3. The first device is the vortex-type separator. The produced fluid, which has already undergone some natural annular separation, is drawn into the unit through the intake ports. These can be straight intake ports, as already mentioned, or a reverse-flow-intake style. The fluid is then boosted to the vortex generator by the positive-displacement inducer. The vortex generator is generally an axial-type impeller. It imparts a high-velocity rotation to the fluid. This causes the heavier fluids (liquids) to be slung to the outer area of the flow passageway and the lighter fluids (free-gas laden) to mingle around the inner area and the shaft. The fluid then enters a stationary flow-crossover piece. The crossover has an outer annular passageway that takes the heavier-density fluids that enter it and directs them to the entrance of the pump. The lighter-density fluid that enters the inner annular passageway of the crossover is directed to the separator vents, where it exits to the casing annulus and flows up the wellbore.

Flanged Connection to Seal-Chamber Section. The bottom flange of the pump bolts to the flange of the seal-chamber-section head. It maintains axial alignment of the shafts of the two units. It also allows the floating pump shaft to engage the end of the seal-chamber-section shaft so that the axial thrust produced by the pump is transferred to the thrust bearing in the seal-chamber section.

Stages. The stages of the pump are the components that impart a pressure rise to the fluid. The stage is made up of a rotating impeller and stationary diffuser. The stages are stacked in series to incrementally increase the pressure to that calculated for the desired flow rate. A graphic of the fluid flow path is illustrated in Fig. 13.4. The fluid flows into the impeller eye area and energy, in the form of velocity, is imparted to it as it is centrifuged radially outward in the impeller passageway. Once it exits the impeller, the fluid makes a turn and enters the diffuser passageway. As it passes through this passageway, the fluid is diffused, or the velocity is converted to a pressure. It then repeats the process upon entering the next impeller and diffuser set. This process continues until the fluid passes through all stages, and the design discharge pressure is reached. This pressure rise is often referred to as the total developed head (TDH) of the pump.

A key feature for both styles of stages is the method by which they carry their produced axial thrust. Usually, the pumps that are under a 6-in. diameter are built as "floater" stages. On these, the impellers are allowed to move axially on the pump shaft between the diffusers. Contrary to the name given to this configuration, the impellers never truly float. They typically run in a downthrust position, and at high flow rates, they may move into upthrust. To carry this thrust, each impeller has synthetic pads or washers that are mounted to the lower and upper surfaces, as shown in the previous figures. These washers transfer the thrust load from the impeller through a liquid film to the smooth thrust pad of the stationary diffuser.

On 6-in. and larger pumps and on specially built smaller pumps, the impellers are usually fixed or locked to the shaft. These pumps are referred to as "fixed impeller" or "compression" pumps. In this configuration, all the thrust is transferred to the shaft and not to the diffuser. Therefore, the seal thrust bearing carries the load of all the impellers plus the shaft thrust. Particular care should be exercised in selecting the proper seal thrust bearing to match the fixed impeller pump conditions because these loads can be very high.

Performance Characteristics. The manufacturers state the performance of their pump stages on the basis one stage, 1.0 specific gravity (SG) water at 60- or 50-Hz power. A typical performance curve for a 4-in.-diameter radial-style pump, with a nominal best-efficiency performance flow of 650 B/D, is shown in Fig. 13.13. A mixed-flow style with a nominal flow rate of 6,000 B/D is shown in Fig. 13.14. In these graphs, the head, brake horsepower (BHP), and efficiency of the stage are plotted against flow rate on the x -axis. Head, flow rate, and BHP are based on test data, and efficiency is calculated on the basis of

The head/flow curve shows the head or lift, measured in feet or meters, which can be produced by one stage. Because head is independent of the fluid SG, the pump produces the same head on all fluids, except those that are viscous or have free gas entrained. If the lift is presented in terms of pressure, there will be a specific curve for each fluid, dependent upon its SG.

The dark (highlighted) area on the curve is the manufacturers recommended "operating range." It shows the range in which the pump can be reliably operated. The left edge of the area is the minimum operating point, and the right edge is the maximum operating point. The best efficiency point (BEP) is between these two points, and it is where the efficiency curve peaks. The shape of the head/flow curve and the thrust characteristic curve of that particular stage determines the minimum and maximum points. The minimum point is usually located where the head curve is still rising, prior to its flattening or dropping off and at an acceptable downthrust value for the thrust washer load-carrying capabilities. The location of the maximum point is based on maintaining the impeller at a performance balance based on consideration of the thrust value, head produced, and acceptable efficiency.

API RP11S2 covers the acceptance testing of ESP pumps. H) is a function of diameter (D) to the second power and also of rotating speed (N) to the second power. Flow (Q) is a function of diameter to the third power and also a direct function of rotating speed.

The BHP curve shows the power required to drive the stage. The power is lowest at shutoff or zero flow and increases with flow. The HP also follows the relationship that is given in Eq. 13.4 for different-sized pumps under dynamically similar conditions.

For any particular-diameter-pump series, there is generally an overlap region between the radial and mixed-flow styles. A typical relationship of a family of similar-diameter stages is shown in Fig. 13.15. Notice that each style increases in efficiency as the flow rate increases, until the efficiency peaks and begins dropping off.

The component located below the lowest pump section and directly above the motor, in a standard ESP configuration, is the seal-chamber section (Fig. 13.16). API RP11S7 gives a detailed description of the design and functioning of typical seal-chamber sections. RP11S7. The seal-chamber section is basically a set of protection chambers connected in series or, in some special cases, in parallel. This component has several functions that are critical to the operation and run-life of the ESP system, and the motor in particular.

Axial Thrust Bearing. This bearing carries all of the axial thrust produced by the pump and seal-chamber section. Generally, sliding-shoe hydrodynamic types are used for this application because of their robustness and ability to function totally immersed in lubricating fluid. It is composed of two main components: a stationary pad and a rotating flat disk. The stationary part has pads finished to a very close flatness tolerance, connected to a base by a thin pedestal or flexible joint. The rotating disk is also finished to a very close flatness tolerance. Several different bearing designs are shown in Fig. 13.22. They represent standard-style cast bearings for normal applications and machined bearings for intermediate- and high-load applications.

The shaft has to transmit, from the motor to the pump, the entire torque required by the equipment for its application. This not only includes the stabilized running torque but also the short-term torque spikes caused by unit startup and intermittent pump loads. Because the diameter of the shaft is constrained because of the maximum diameter of the unit, materials of differing mechanical properties must be used to provide different load capabilities. These materials must also provide protection from corrosive wellbore fluids.

The thrust-bearing performance is a function of the load that is transferred to it and the viscosity of its lubricating oil. The load transmitted from the pump can be calculated on the basis of the pump geometry and the TDH produced for the application. For "floater" pumps, the shaft load is always down and is equal to the cross-sectional area of the top of the shaft multiplied by the discharge pressure of the pump (Pdischarge) minus the cross-sectional area of the bottom of the shaft multiplied by the pump intake pressure (PIP). For "fixed" impeller pumps, the load is equal to the shaft force, as just calculated, plus the summation of all the impeller thrust forces. The impeller thrust forces can be roughly calculated, as previously described in the pump-stage section, or obtained from the pump manufacturer.

Revolutions per Minute (RPM). The rotational speed or RPM of the motor at its application load point is very important in determining the operating point or output of the pump. The pump-performance curve used in determining the head and flow output of the pump for its application is based on a pump-motor speed of 3,500 RPM. If the RPM varies from 3,500, the pump flow will vary with the ratio of the speed, and the flow rate will vary with the ratio of the speed squared. (See Eqs. 13.1 and 13.2.) Once again, by knowing the percent of nameplate amps, the motor speed can be read from the motor characteristic curve. Even though this RPM change is usually small, it can still impact the final motor and pump operating point for a particular application. When the pump-performance point is modified, because of the motor RPM, the pump head and flow rate change; therefore, the load on the motor is changed. Determining the final pump operating point and motor loading point becomes an iterative process.

Motor Lead Extension (MLE). The motor lead extension cable, also referred to as the motor flat, is a specially constructed, low-profile, flat cable. It is spliced to the lower end of the round or flat main power cable, banded to the side of the ESP pump and seal-chamber section, and has the male termination for plugging or splicing into the motor electrical connection. Because of its need for low profile, it requires compact construction. It generally has a thin layer of high-dielectric-strength polyamide material wrapped or bonded directly to the copper conductors. This allows for a thinner layer of insulation material, allowing for a lower profile. The MLE is generally selected on the basis of equipment: casing clearance and the voltage capacity requirement.

Control Module. These are solid-state devices that offer basic functions necessary to monitor and operate the ESP in a reliable manner. The unit examines the inputs from the CPT and other input signals and compares them with preprogrammed parameters entered by the operator. Some of the functions include overload and time-delayed underload protection, restart time delay, and protection for voltage or current imbalance. Additional external devices can be connected, which provide for downhole pump intake pressure protection, downhole motor temperature protection, surface tank high/low level controls, line pressure switches, and others.

VSCs used with ESPs should be designed for the specific requirements of the downhole ESP motor and pump. This is because of the unique design and characteristics of the downhole centrifugal pump and submersible motor as compared to their surface counterparts. Generally, the VSC is designed to provide a constant volts/hertz output through a broad range of frequency variations. The magnetic flux that is generated in the stator of the submersible motor and passes through the rotors is directly proportional to the voltage and inversely proportional to the frequency of the applied power. The result is a constant magnetic flux density in the motor. Because the output torque of the motor is proportional to the magnetic flux density, the motor is a constant-torque variable-speed device. Also, because of its low inertia characteristics and unique rotor design, it does not have the same high-operating-speed restrictions as a typical surface induction motor. Therefore, a VSC is typically applied to frequencies from 30 to 90 Hz, with its minimum and maximum frequencies restricted only by the mechanical limitations of the downhole ESP equipment.

Because of the relationship of the performance of a centrifugal pump to its rotational speed (Eqs. 13.2 through 13.4), the VSC allows for wider flexibility of the downhole ESP system. The effect on pump operation is shown in Fig. 13.34. This is the same pump that is represented in the 60-Hz fixed-speed performance curve of Fig. 13.14. This allows the designer to select the flow rate and speed of the system on the initial design. For this pump stage, it can be operated between 1,800 B/D at 30 Hz (minimum recommended operating point) and 10,200 B/D at 90 Hz (maximum recommended operating point). The benefits of VSC usage are discussed next.

Broadened Application Range. On fixed-speed operation, a pump stage has a recommended minimum and maximum flow rate. Beyond these points, the pump can operate in a detrimental run-life or reliability area. By operating at reduced frequency, the minimum recommended operating point is reduced, and, at higher frequencies, the maximum operating point is increased. This allows the application of ESPs in low-productivity-index (PI) wells and higher flow rates to be obtained from small bore casings. It also allows a limited inventory of pumps to be applied over a broader flow range.

Maximize Well Production. If the well PI is greater than that for the original design, either through data error or changing wellbore parameters, the ESP operating point can be increased with a VSC. The HP rating of the motor limits the frequency increase. Remember, the HP load from the pump increases with the cube of the frequency ratio, and the HP capability of the motor increases directly to the speed ratio. Therefore, the designer must consider using an oversized motor if there is a potential need of higher flow rates.

Minimum Well Production. If the well PI is lower than that for the original design, the ESP operating point can be decreased with the VSC. The TDH of the pump is the limiting factor on the minimum VSC frequency. The produced head of the pump decreases with the square of the frequency ratio. Therefore, the designer must consider initially oversizing the pump lift, if there is a potential for reduced-frequency operation.

Pump Intake or Casing Annulus Pressure. This information provides wellbore static pressure and the well flowing pressure at the production rate. If the measurement is sensitive enough, it can also provide excellent well drawdown information.

Pump Discharge Pressure. This parameter provides a reading on the discharge pressure of the pump. This reading and the pump intake pressure provide a measurement of the TDH of the pump. Comparing this value to the design TDH, hydraulic performance of the pump can be monitored and continually evaluated. Additionally, for gassy and/or viscous fluids, pump-performance correction factors can be established or verified for that particular wellbore condition.

Pump Discharge Temperature. This measurement provides the temperature of the discharge fluid from the pump. The production fluid is heated as a result of the heat rejected by the motor and pump inefficiencies. The fluid heat rise through the pump can be used to calculate the fluid volumetric increase and the viscosity change of the fluid. Once again, sudden spikes or longer-term changes can provide warnings of potential problems.

Downhole Flow Rate. Downhole flowmeters are available that provide flow-rate measurements from the pump discharge. This is an excellent tool, when compared to the surface flow rate, for evaluating ESP performance and warning of potential problems. Because surface flow rate is not generally continuously monitored, this can be a piece of information for enhanced ESP protection. In multiphase-fluid (gassy) applications, the selection and calibration of the flowmeter is important because of the difficulty in accurately measuring this fluid.

Packers are used with ESP systems when there is a need to isolate the annular area above the ESP and/or provide a positive barrier between the pressurized wellbore fluid and the area above the packer. Isolating the area above the packer is done to segregate two separate zones or prevent or reduce the rate of wellbore fluid corrosion damage to the casing. With a deep-set packer, operational precautions must be observed to prevent damage to the ESP system. With a deep-set packer, the volume contained between the packer and pump intake is usually small. Upon startup, the ESP can evacuate this volume quickly, causing a sudden drop from wellbore static to flowing pressure. This causes sudden decompression to the cable and internal volumes of the seal-chamber section and motor, especially if they have been saturated with solution gas. This decompression can cause expansion and insulation damage to the cable. If it is severe enough, it can result in extensive expulsion of motor oil from the seal-chamber section and motor, possibly rupturing elastomer seals and bags.

Centralizers/Protectorilizers. Centralizers are sometimes used when the ESP is installed in a deviated wellbore or into a tapered-string casing. Its function, when used in a deviated wellbore, is to be a contact point with the casing and allow the ESP unit to have some standoff clearance. They are typically located at the bottom of the ESP unit and, in some cases, at points along its length or at the discharge tubing. They have to be constructed so as not to restrict the flow by the motor and to the pump intake. Generally, they are designed with at least three radial fins attached to either tubing, for the top and bottom unit or to metal straps, which can be attached around the ESP body. Centralizers are also used when an ESP is deployed into a tapered-string casing. Its function is to help guide the unit into and through the casing step to reduce the chance of mechanical damage. It is normally a finned configuration with the bottom end tapered or bull nosed.

Check/Drain Tubing Valves. A check valve is used in the production tubing string, generally two to three joints above the pump discharge, to maintain a full column of fluid above the pump. This may be desired to eliminate the time it takes to raise the fluid from its static fluid level to the surface ("pump-up time") or the protective shutdown time for fluid fallback. Normally, each time an ESP cycles off, the fluid falls back from the surface to its static fluid level. On restart, it again has to lift the fluid from its static point to the surface. Holding the fluid in the tubing can eliminate this. Also, when the fluid is falling back, it causes the de-energized pump to spin backwards. If power is applied during this period, damage to the ESP could result. Generally, a backspin sensor or restart timer is used on the motor controller for premature restart protection.

The use of a check valve should be reviewed in gassy or high-GOR wells and wells that produce significant solids. In a gassy well, when the unit shuts down, a gas cap can form under the check valve and be held there by the fluid column above the check. If the gas cap volume is large enough to extend down to or below the pump intake, the pump will be immediately gas locked and unable to pick up a prime. When there are solids (especially sand) entrained in the production fluid and the ESP is shut down, the solids fall back in the production tubing and settle either on the check valve or into the pump discharge. This could either plug the tubing above the check valve or the pump. Therefore, the use of a check valve in fluids with solids should be reviewed.

Motor Shroud/Recirculation Systems. Shrouds, as shown in Fig. 13.36, are used to redirect the flow of production fluid around the ESP system. The shroud assembly is made up of a jacket (a length of casing or pipe), a hanging clamp and sealing retainer for the top, and a centralizer for the bottom. The jacket dimensions are selected on the basis of shroud location relative to the production source and the function of the shroud. But, at a minimum, the shroud should extend to below the bottom of the motor. The shroud ID has to allow for the insertion of the ESP with flow clearance to allow for proper cooling velocities without choking or excessive pressure drop to the flow. The shroud OD must have sufficient clearance with the casing ID to assure reliable deployment and proper flow from the well perforations to the pump intake. Fluid pressure drop in this annular area, similar to the shroud-to-ESP annular area, can be significant enough to impact the pump intake conditions.

The most commonly used shroud configuration is shown in the left graphic of the same figure. In this configuration, the ESP is set below perforations and the shroud directs the production flow down and back up by the motor for cooling. Otherwise, the fluid would be pulled down to the pump intake, leaving the motor in stagnate fluid with heat rise concerns. The purpose of setting below perforations is to increase the production rate for the same pump intake pressure or to serve as a simple reverse-flow gas-separation system. In the gas-separation application, the configuration depends on the free gas flowing from the perforations taking the path of least resistance—up the open casing annulus, instead of down to the bottom of the shroud. One caution, in this configuration, is not to use a gas-separation intake on the pump. The vented free gas from the separation intake would recycle to the bottom of the shroud, increasing the free-gas ratio to the pump and decreasing the cooling of the motor.

This configuration is not recommended for setting above perforations in an application with free gas. But where the ESP and casing annular area is large, creating too low a cooling flow, a shroud can be used to increase the production-fluid cooling velocity. For those special cases of setting above perforations and the problem of free gas, an inverted shroud (right graphic in the figure) has proved successful in separating free gas from the fluid that is directed back down to the pump intake.

In wells that have a diameter restriction because of tapered casing, liners, or screens, a stinger can be attached to the bottom of the shroud to position the intake below perforations and down into the restriction. A stinger is a section of tubing, usually smaller in diameter than the shroud, which is attached to the bottom of the shroud and provides fluid communication from the wellbore to the interior of the shroud. This configuration is shown in the center graphic of the figure. The pressure drop through the stinger must be calculated to check for possible choking of the pump and also for an increase in the free gas liberated, causing gas interference issues with the pump and cooling issues with the motor.

Screens and Filters. Screens and filters are used with ESP systems to prohibit the flow of large solids into the pump intake. In one configuration (shown on the intake of Fig. 13.2), a mesh screen or perforated metal sheet is wrapped or mounted over the pump intake ports. The mesh or perforation size has to be small enough not to allow the passage of large particles, but large enough not to cause a flow restriction. The size of particle that must be screened is a function of the flow-passageway clearances through the pump. If a shroud is used, a screen can be used to cover the open intake area at the bottom of the shroud.

Filters have also been used on ESP applications. The simplest method is to use a motor shroud with a stinger, shown in the center graphic of Fig. 13.36. The stinger is sealed at the end, perorated along its length, and a filter element or gravel pack is inserted into or around the stinger. The production fluid then has to pass through the stinger filter prior to entering the pump intake.

Several cautions must be mentioned if screens or filters are used. The open area of the screen must be several times larger than that of the open area of the pump intake ports. This allows for proper flow without choking when, not if, the screen starts building up debris and plugging. This is also the case with the filters. Also, remember that the separated debris has to go somewhere and that is generally in the rathole below the ESP. The rathole must be large enough to hold the amount of debris expected over a period of time. This is because if it starts building up on the ESP, it can cause motor heat problems, eventual complete plugging of the intake ports, and difficulty in pulling the unit. Plugged screens and filters may cause severe pump and motor problems, if not designed and applied correctly.

What has been described up to this point is the standard ESP configuration. It has the pump, seal-chamber section, and motor attached to the production tubing, in this order from top down. In some wellbore completions and unique ESP applications, the arrangement and configuration of the system is modified. Some of these applications are listed next.

Inverted Bottom-Intake ESP. An inverted-unit configuration has the motor on top, attached to the tubing string; seal-chamber section underneath the motor; and the pump on bottom (Fig. 13.38). For a bottom-intake design, the production fluid is drawn in the intake ports located at the very bottom of the ESP system and discharged out of ports located just below the connection to the seal-chamber section. Because the discharged production fluid cannot flow through the seal-chamber section and motor, it has to exit into the casing or liner annulus and flow past these units. Once above the motor, it can continue flowing up the annulus or be ported back into the production tubing string. Additionally, the casing annulus communication flow path, between the intake and discharge ports, has to be sealed to prevent recirculation. Generally, the intake is stung into the casing packer to seal this path. This configuration is typically used for applications in which the intake needs to be located as low as possible, cavern or mine applications, annular flow designs, coiled tubing with internal power cable, or cable-deployed ESP systems.

Inverted Bottom-Discharge ESP. This design is configured the same as the inverted bottom-intake ESP system with the exception that the pump stages are inverted to pump down (Fig. 13.39). Once again, the intake and discharge fluid communication path in the casing annulus has to be closed. Generally, the pump discharge, on the bottom of the ESP assembly, is stung into an isolation packer. The wellbore production fluid is transferred from above this packer to below under high enough pressure to inject into the lower formation. This configuration is typically used for injection of water into a disposal zone.

Special designs that incorporate downhole hydrocyclone separators have been used to separate some of the water from the wellbore fluid (Fig. 13.40). In this case, the reduced-water-content oil is pumped to the surface, and a significant portion of the deoiled water is injected into a disposal zone.

Series Production. A dual-ESP system can also be used for high total-developed-head requirements. This is where the lift requirement or pressure increase across the pump is beyond the equipment design limitations. By connecting the ESP systems in series, large pressure increases can be achieved for the desired flow rate while staying within each individual unit’s HP and burst-pressure limitations (Fig. 13.42).

Booster ESP. The ESP can also be used as a pressure boost system for surface applications. They can handle a wide variety of fluid conditions and do not have the pressure pulsation attribute associated with positive-displacement-type pumps.

Surface Horizontal System. This configuration utilizes an ESP centrifugal pump driven by a surface electric motor, engine drive, or other primary mover. It is generally mounted on a skid for stability and alignment (Fig. 13.44). It can provide a nonpulsating flow and a wide flow range with the use of a variable-speed drive.

Pipeline-Insert System. In this configuration, the ESP is inserted into a parallel section of piping. Fluid can then either flow directly through the pipeline or can be valved to bypass through the pump leg section for pressure boosting.

Through-Tubing-Conveyed ESP. In applications where pump wear and intervention costs are a major concern, a through-tubing-deployed pump is an option. The configuration is shown in Fig. 13.45. The motor and seal-chamber section are deployed on the bottom of a tubing string. The power cable is connected to the motor and deployed with the tubing, locating and protecting it in the casing/tubing annulus. The pump section is then deployed by a work string, typically wireline or coiled tubing, and latched onto the seal-chamber section. Thereafter, workovers, because of pump issues, can be done at a lower expense with wireline or coiled-tubing rigs, instead of regular jointed-tubing workover rigs.

Multiphase Flow.Performance Variables. The amount of free gas that an ESP pump can handle is a function of the following variables: pump-stage geometry, operating point of the pump stage, control by a fixed-speed or variable-speed drive, pump-intake flowing pressure, and wellbore geometry.

Pump-Stage Geometry. The gas handling capability of a centrifugal pump stage increases with flow rate or stage specific speed—a nondimensional design parameter. In other words, as the stage style moves from radial to mixed flow (Fig. 13.11), the gas-handling capability increases.

Pump Operating Point. The most stable operating region for a pump stage on gassy fluid is from the maximum recommended flow rate back to its BEP. As the flow rate moves from the BEP toward the minimum recommended operating point, the potential for gas interference affecting pump performance is increased.

VSC Operation. The VSC allows for some additional flexibility and reduction in unit shutdowns that are related to pump gas locking. Tests have shown that the pump gas-handling capability increases slightly with increasing speed. If the pump load decreases and the motor amps drop, indicating an initiation of gas lock, the VSC can be programmed to speed up for a short period to attempt to clear the gas-lock situation. If it clears and the load picks back up, the VSC would then return to its set operating frequency. If it does not clear, the unit would then shut down on an underload situation and restart on the time out delay.

Pump-Intake Pressure. The gas handling capability of the pump is very sensitive to pump-intake pressure. An empirical correlationFig. 13.46. The area under the curve represents stable operation, and the area above indicates potential gas-interference and -locking regions.

Tapered Pumps. Tapered pumps utilize several different sets of pump stages in the same pump housing or pump string. Generally, the first section of stages is mixed-flow style because they can handle a higher percentage of free gas. As the gassy fluid is pressurized through each of these first stages, the total fluid volume decreases because of the compression of the free gas. When the flow rate nears the BEP flow rate of these stages, a second set of stages is selected. Generally, a good design can be accomplished with two or three sets of stages in the taper.

Mechanical Separation. The vortex and rotary separation intake components, which were discussed in the pump section, are used here to add centrifugal separation to the gassy fluid that enters the intake section. Because there are so many variables that affect their effectiveness or efficiency, the manufacturers should be contacted for separation efficiency values or guidelines. These units can also be used in tandem to accomplish series separation.

Recirculation Pump. In a completion scheme where there is insufficient clearance to run a shrouded unit below perforations, a recirculation pump can be used. A recirculation pump bleeds a small portion of the pumped fluid off and circulates it down below the motor by a small-diameter hydraulic tube. This establishes a small flow in the rathole where the motor is set. By properly designing the bleed flow, cooling flow by the motor can be maintained. Since the perforations are above the unit and pump intake, natural annular gas separation can be maximized.

Abrasive Slurries.Performance Impact of Abrasives. There are three types of wear that impact the pump stage and its performance. They are listed next and prioritized in order of importance or impact.

Radial Wear. As the slurry wears the radial-support bushing system of the pump, it loses its lateral stability. This allows the rotating parts to start interfering with the stationary parts. Vibration increases, and it starts impacting the top of the seal section where the first mechanical shaft face seal is located. Once vibration and radial movement start to influence the face seal, leakage starts across the sealing face. This initiates a path for the well fluid to progress toward the motor.

Downthrust Wear. On the floating-style stages, the abrasive slurry migrates into the downthrust bearing pad area of the pump stage. The stationary diffuser thrust pad starts boring into the impeller thrust washer area. Once it breaks through the lower shroud of the impeller, the impeller loses part of its work to recirculation flow. As the diffuser pad bores further into the impeller passageway, it also blocks a portion of the impeller flow path, thus restricting the remaining flow.

Erosion Wear. As with any abrasive-slurry flow along a twisting path, erosion wear takes place. Although it is not usually associated with the failure of the pump, it is a potential failure mode and a concern, especially when modifications have been made to the pump to address the radial and downthrust wear modes. Erosion wear not only damages the stage pieces, it also wears any surface with which it comes into contact. Severe cases have resulted in the wear perforating the pump or production-tubing walls and dropping units in the well.

Compression Pumps. For many years, this was the answer for abrasive applications. In a compression or "fixed-impeller" pump, the impellers are fixed to the shaft or stacked hub to hub so there is no axial movement. With all the impellers fixed relative to the shaft, the whole impeller stack can be raised slightly so that it does not run into contact with the downthrust or upthrust pads on the diffuser. This pump design eliminates the downthrust wear mode. When it is used in conjunction with hardened journal bearings, it also addresses radial wear problems. There are several issues with compression pumps. First, they are very difficult to assemble properly. Because an ESP pump is a very long, multistaged assembly, it is very difficult to locate all of the impellers and still have the needed minimum shaft axial movement. Also, now, the thrust of each impeller is transferred to the shaft and is added to the normal shaft thrust produced by the discharge pressure on the top area of the shaft. The thrust bearing in the seal-chamber section is required to carry this additional thrust. Additionally, as the sealing areas of the pump stage wear, the downthrust also increases. Therefore, the selection of the proper thrust bearing is critical, and the anticipated thrust must be calculated on the basis of the maximum thrust seen from worn stages.

Thrust and Radial Protection. In this modification, the base material in the radial and downthrust areas of the stage is replaced with inserts of hardened materials. The materials are usually tungsten or silicon carbides, or ceramics. This results in a pump with both radial and downthrust protection but is built in a floater style.

Viscous Crude and Emulsions. ESPs are also used to lift viscous fluids, commonly referred to as heavy and extra-heavy crudes. Viscosity is defined as the resistance of a fluid to movement as a result of internal friction. Resistance causes additional internal losses in a centrifugal pump. The increases in internal losses of a centrifugal pump affect each performance parameter.

Performance Impact of Fluid Viscosity. Effect on Flow Capacity. Flow capacity of a given pump stage diminishes rapidly with a relatively small increase in viscosity. The rate of correction tends to moderate as viscosity continues to increase. The amount of correction is also dependent on stage geometry, and the decrease in capacity is more exaggerated for radial flow stages.

Effect on Head. The total dynamic head at the BEP diminishes on a moderate curve as viscosity increases. It is affected to a lesser extent than flow capacity. The head at zero flow remains relatively constant. Fig. 13.47 shows various head vs. flow-rate curves for an ESP pump stage rated for about 2,100 B/D on water.

There are several published methods for estimating the effect of viscosity on the head, flow rate, and BHP of a centrifugal pump. These "standard" correction factors are usually not accurate for the specific small-diameter, multistage design of ESP pumps. Therefore, most manufacturers have established corrections through testing for each pump stage type in their product line. These correction factors are based on dead-oil viscosity values for the fluid at pump-intake conditions. When applying these corrections to the pump, the following should also be considered.

Effects of Gas. When gas saturates into the crude, it reduces the viscosity of the fluid. Some amount of gas is helpful in reducing fluid viscosity, but an excessive amount of free gas is disruptive to well fluid production. Gas tends to migrate out of highly-viscous fluid slowly. Therefore, a higher percentage of gas tends to pass through the pump with the produced well fluid. In an application with gas, the designer must be aware of two viscosity values. The first is the dead-oil viscosity. This is the viscosity of the crude at dead or completely degassed conditions. The other is the live-oil viscosity. It is the apparent viscosity of the gas-saturated crude and the viscosity that affects the pump performance in a well with gas. There are several dead-oil and saturated-oil viscosity correlations that can be used during the design process. The correlation selection should be based on modeling of the actual wellbore performance.

Effect of Temperature. Temperature has a dramatic effect on the viscosity of the crude oil. Therefore, it is critical to the ESP design process that the fluid temperature in the wellbore at the pump setting depth is known. This determines the fluid viscosity and pump-performance correction factors at the first pump stage. Additionally, the inefficiency of the pump results in additional heat loss to the fluid and surrounding wellbore. This incremental elevation in temperature from stage to stage through the pump moderates the impact of the fluid viscosity on the total pump performance. Therefore, the designer should, at a minimum, use an average viscosity for the fluid through the pump for sizing applications. A more accurate method is to calculate the performance on a stage-by-stage basis, using the fluid input conditions to each stage. Most design software programs use this method.

Dilution. Some success has been achieved with diluent injection. In this process, a lighter crude or refined product, such as diesel, is injected from the surface via a separate hydraulic line to a point below the ESP or directly into the pump intake. This effectively cuts the viscosity of the wellbore fluids. The amount of injected diluent depends on the desired final mixture viscosity. This type of viscosity reduction also reduces the surface flowline losses, which reduces the required wellhead pressure or the need for diluent injection at the wellhead. Using a diluent fluid is an effective, but expensive, approach.

Chemical Injection. Viscosity-reduction and emulsion-breaking chemicals can be injected from the surface by hydraulic injection lines. This impacts the fluids in the annulus and through the pump but not very far back into the reservoir.

Another source of corrosion is hydrogen sulfide, H2S. The H2S mainly attacks copper-based alloys of the pump, seal-chamber section, and cable. This type of corrosion can be controlled by replacing the copper-based alloy components with suitable materials or by isolating them from exposure to the well fluid and gases. When CO2, H2S, and hot brine are combined, unpredictable corrosion results may appear. With small changes in the concentration of CO2 and/or H2S and temperature, corrosion could even vary significantly from well to well within the same reservoir.

Another corrosion mechanism that has been around the oil field for years, but has been misunderstood or misdiagnosed, is microbiologically influenced or induced corrosion. Scale and Asphaltenes. If the well has scaling or asphaltene-forming tendencies, these can be detrimental to the performance and run-life of the total ESP system. Because of the characteristics of the ESP system, there are pressure and temperature changes, which provide a mechanism for scales to form or precipitate out of solution. Typically, scales cause two problems. They plug the flow passageways of the pump stages, reducing or stopping the flow entirely. They also adhere to the outside surfaces of the motor and seal-chamber section, reducing the heat-transfer rate, causing both units to run hotter. Asphaltenes generally only cause plugging of the pump stages. Both problems can be reduced, but not totally eliminated, by applying synthetic coatings to the surfaces affected or by using a downhole inhibitor-chemical treatment.

Production of high-viscosity fluids can result in significant flow losses through the production tubing and surface piping. In some instances, the pressure requirements generated because of flow losses can exceed the hydrostatic head on a well. As discussed previously, pressure losses in the system accumulate and are reacted at the pump, where they cause additional pump pressure loading, leading directly to increased rod-string axial loads and system torque. It is critical that system design account for the "worst-case" flow losses, particularly the selection of the pump (pressure rating), rod string (torque capacity), and prime mover (power output).

Fig. 15.35 shows a good example of the effects that viscous flow losses and water slugging can have on pump loads in a heavy oil well. The axial and torsional loads on the well were monitored in real time with a purpose-built PCP system dynamometer unit. Fig. 15.35. Because the only significant difference during the operating period was the viscosity of the fluid being pro